Hybrid cascade refrigeration systems for refrigeration and heating

Abstract

In this study a cooling ejector cycle coupled to a compression heat pump is analyzed for simultaneous cooling and heating applications. In this work, the influence of the thermodynamic parameters and fluid nature on the performances of the hybrid system is studied. The results obtained show that this system presents interesting performances. The comparison of the system performances with hydrofluorocarbons (HFC) and natural fluids is made. The theoretical results show that the a low temperature refrigerant R32 gives the best performance.

Full text

archives
of thermodynamics
Vol. 39(2018), No. 2, 73–95
DOI: 10.1515/aoter-2018-0013
Hybrid cascade refrigeration systems for
refrigeration and heating
K. MEGDOULI1,2∗
Y. EZZAALOUNI1
E. NAHDI1
A. MHIMID2
L. KAIROUANI1
1Unité de Recherche Energétique et Environnent, Ecole National
d’ingénieur de Tunis 37 Le Belvédčre Tunis, Tunisie.
2Laboratoire d’Études des Systèmes Thermiques et Énergétiques Ecole
National d’ingénieur de Monastir Avenue Ibn El Jazzar, 5019 Monastir,
Tunisie
Abstract In this study a cooling ejector cycle coupled to a compression
heat pump is analyzed for simultaneous cooling and heating applications. In
this work, the influence of the thermodynamic parameters and fluid nature
on the performances of the hybrid system is studied. The results obtained
show that this system presents interesting performances. The comparison of
the system performances with hydrofluorocarbons (HFC) and natural fluids
is made. The theoretical results show that the a low temperature refrigerant
R32 gives the best performance.
Keywords: Ejector cycle; Heat pump; Comparison; Refrigerant; Coefficient of perfor-
mance; Irreversibility; Exergy loss; Second law efficiency
Nomenclature
COP – coefficient of performance
Cp– specific heat of gas at constant pressure, kJ/kgK
c– compressor
∗Corresponding Author. Email: karima.megdouli@gmail.com

74 K. Megdouli, Y. Ezzaalouni, E. Nahdi, A. Mhimid and L. Kairouani
D– diameter of the mixing chamber, m
d– diameter the primary nozzle exit, m
dD– diameter at the diffuser exit, m
d∗– diameter of the primary nozzle throat, m
F– coefficient of wall friction
fx– friction force per unit of mass, N/kg
H– efficiency or performance
h– specific enthalpy, kJ/kg
I– irreversibility, KW
K– ratio of specific heats (=Cp/Cv)
L– length of the mixing chamber, m
M– Mach number
m– mass flow rate, kg/s
Q– heat load rate, kW
P– pressure, Pa
R– ideal gas constant (J/kgK)
r– compression ratio
S– cross-section area, m2
S– specific entropy, kJ/kgK
T– temperature, K
T0– environmental temperature, K
Tb– temperature of the boiler
Tc– temperature of the condenser
Tg– temperature of secondary fluid in the boiler, K
Th– temperature of secondary fluid in the condenser, K
Tr– temperature of secondary fluid in the evaporator, K
U– entrainment ratio of the ejector
X– position of the primary nozzle relative to the secondary nozzle throat, m
˙wp– pump power
˙wc– compressor power
U– entrainment ratio
V– velocity, m/s
˙
W– mechanical work rate, kW
Greek symbols
ηII – second law efficiency, %
φ– area ratio between mixing tube and primary nozzle throat
(= (D/dx)2)
ξ– driving pressure ratio
∆– related to the variation of a parameter
ρ– density, kg/m3
η– compressor efficiency, %
Exponents and Subscripts
c– compressor
e– evaporator

Hybrid cascade refrigeration systems for refrigeration and heating 75
B– boiler
Ev – expansion valve
Int – intercooler
cd – condenser
Ej – ejector
D– diffuser
p– pump
* – primary nozzle throat
′– primary fluid (or motive fluid)
′′ – secondary fluid (or entrained aspirated)
1 Introduction
Many industrial sectors (food, chemical, and pharmaceutical) require low
temperature cooling and high temperature heating simultaneously which
cannot be achieved simultaneously and effectively by single stage or multi-
stage systems due to individual limitation of refrigerant. These needs can
be achieved by using a cascade system.
Bhattacharyya et al. [1] examined CO2-C3H8cascade system for simul-
taneous refrigeration at −40 ◦C and heating at 80 ◦C. A parametric study
on NH3-CO2cascade refrigeration system was reported for a temperature
range of −50 to 40 ◦C [2]. Bhattacharyya et al. [3] analyzed a natural
refrigerant based cascade system with nitrous oxide as the low temperature
fluid and carbon dioxide as the high temperature fluid for simultaneous
cooling and heating applications. They found that for a given set of oper-
ating parameters the optimized intermediate temperature for a maximum
coefficient of performance (COP) varies with the approach temperatures
of the gas cooler and the evaporator in such a way that the nitrous ox-
ide condensing temperature in cascaded heat exchanger remains the same.
They found also that the system performance is independent of fluids used
in high temperature and low temperature cycles, since N2O and CO2have
very similar thermodynamic properties.
Using low temperature energy source applications such as solar energy,
geothermal energy and waste heat for simultaneous refrigeration and heat-
ing applications is the primary objective of the present study. This paper
describes a hybrid system that combines a basic compression heat pump cy-
cle with an ejector cooling cycle. The interface between both systems is an
intercooler. A schematic diagram of the combined system is shown in Fig. 1.
The pressure-enthalpy diagram is presented in Fig. 2. For thermodynamic
reasons, the working fluid exiting the compressor providing a considerable

76 K. Megdouli, Y. Ezzaalouni, E. Nahdi, A. Mhimid and L. Kairouani
thermal energy and exergy which will be dissipated if not used. This is an
ideal thermal energy source to be used in heating application or for power
generation. To our knowledge, this sort of waste heat utilization has not
been investigated yet and worth to be paid attention. In particular, there
is a lack of knowledge when this heat utilization is performed in a cascade
system with an ejector.
Figure 1: Schematic of the cascade cycle for refrigeration and heating.
Many hydrofluorocarbon (HFC) refrigerants and natural fluids will be an-
alyzed in this cycle to find the most suitable fluid that allows the highest
performance. To ameliorate this system we analyze thermodynamic models
with ejectors. We will use a one dimensional based on the constant area
mixing ejector at critical mode. To locate sources of losses, we will investi-
gate the behavior of the cascade system using the exergy method.
The refrigeration and heat pump cascaded system has been modeled
employing energy conservation on each individual component of the sys-
tem. The following assumptions have been made to simplify the analysis:
1. Heat transfer in heat exchanger with the ambient is negligible.
2. The compression process in the compressor is adiabatic and irre-
versible.
3. The expansion process is isenthalpic.
4. Pressure drop in the connecting pipes and heat exchangers are negli-
gible.

Hybrid cascade refrigeration systems for refrigeration and heating 77
Figure 2: Pressure-enthalpy (p-h) diagram for the cascade cycle for refrigeration and
heating.
2 Energetic analysis
The ejector entrainment ratio, U=m1/m2is defined as the ratio of the
ejector suction mass flow rate at state (1) to the motive mass flow rate at
state (2). For a mass flow rate of 1 kgs−1refrigerant mixture in the ejector,
the suction and motive mass flow rate will be U/(U+ 1) kg s−1and 1/(1 +
U) kgs−1, respectively. The compressors consumption powers, the heat
load rate and the absorbed heat are defined by the following expressions:
˙
Wcomp = ˙m1(h8−h7),(1)
where
h8=h7+(h8is −h7)
ηs
, h8is =h8is (P=P9, s =s7)
The compressor efficiency ηsis given by Brunin et al. [11] as
ηs= 0.874 −0.0135P8
P7
,

78 K. Megdouli, Y. Ezzaalouni, E. Nahdi, A. Mhimid and L. Kairouani
˙
Wp= (h6−h4)1
1 + U,(2)
˙
Qe= (h1−h5)U
1 + U,(3)
˙
Qb= (h2−h6)1
1 + U,(4)
˙
Qcd = ˙m1(h8−h9).(5)
The energy balance at the cascade heat exchange is expressed by
h3−h4= ˙m1(h7−h10),(6)
˙m1=h3−h4
h7−h10
.(7)
The system performance could be evaluated using its coefficient of perfor-
mance, which is defined as
COP =˙
Qe+˙
Qcd
˙
Qb+˙
Wcomp +˙
Wp
.(8)
3 Ejector modeling
The ejector consists of two nozzles: a primary nozzle also called the driv-
ing nozzle, secondary nozzle, mixing tube and diffuser [4,5]. In the driving
nozzle, the enthalpy of the fluid is converted into kinetic energy. At the
output of the driving nozzle, the jet velocity is very high, causing a negative
pressure around the outlet section.
At the same time, due to a strong interaction between the two jets,
resulting in exchange of mass and momentum, fluid from the secondary
nozzle is sucked and brought by the primary fluid, producing entrainment
phenomenon. The mixing process takes place in the mixing chamber, pro-
ducing a kinetic energy transformation of the jet engine into enthalpy of the
mixture and therefore by an increase in the pressure of the latter. At the
outlet of the mixing chamber, the fluids sufficiently mixed pass through the
diffuser, in which there will be a conversion of kinetic energy into pressure.
Le Grives and Fabri defined three different operating regimes for ejec-
tors [6]. In these definitions, the ejector operating regimes are named based
on the dependence of the entrainment ratio on the back pressure at the exit

Hybrid cascade refrigeration systems for refrigeration and heating 79
of the ejector (or the driving pressure ratio). These regimes are the super-
sonic regime (SR), the transition regime (TR) and the mixed regime (MR).
The performance analysis of an ejector consists of determining the forma-
tion conditions of these regimes. During the operation of ejector in the SR,
since the primary static pressure at section e1shown in Fig. 3, is higher
than that of the secondary vapor, the primary fluid expands against the
secondary fluid and causes the velocity of the secondary fluid to reach su-
personic speed at the aerodynamic throat formed by it. As a consequence of
this secondary stream choking phenomenon, the secondary mass flow rate
becomes independent of the back pressure. The MR includes all the cases
for which the secondary flow is not choked. The secondary flow cannot
reach sonic speed within the mixing chamber, and therefore, its mass flow
rate changes depending on the back pressure. In the TR, the secondary
vapor reaches supersonic speed at the point of confluence of the primary
and secondary vapours. It gives the optimum performance of the ejector
[7]. For this reason, we will analyze the ejector in TR. The ejector mod-
eling in transition regime developed by [8] by using the one-dimensional
constant-area ejector model is modified so that it could be applied to the
ejector-compression cycle for cogeneration of heat and cold.
Figure 3: Schematic of a constant-area ejector and geometric parameters [9].
The following assumptions are made for the analysis:
1. The flow inside the ejector is considered one-dimensional homoge-
neous equilibrium flow.
2. The working fluid is an ideal gas with constant properties Cpand k.

80 K. Megdouli, Y. Ezzaalouni, E. Nahdi, A. Mhimid and L. Kairouani
3. For simplicity in deriving the 1D model, the isentropic relations are
used before mixing.
4. The two fluids are completely mixed at the exit of the mixing cham-
ber.
5. X6= 0.
6. Kinetic energies of the refrigerant at the ejector inlet and outlet are
negligible.
With X6= 0 the flow in convergent part is sonic which implies that the
aerodynamic throat is situated in the cylindrical part of the mixing cham-
ber. So, we deduce that, for TR with a distance X6= 0, the aerodynamic
throat is located at the entry of the mixing chamber. Therefore
M′′
e2= 1 .(9)
To form the sonic throat of the secondary fluid at the section e2, the motive
flow must expand, which imposes P′
e2> P ′′
e2. After the section e2, we can
only have P′
e2> P ′′
e2, since the case P′
e2< P ′′
e2, is physically impossible. For
the condition P′
e2> P ′′
e2, the primary fluid is going to continue to expand,
the sonic throat is situated then downstream the section e3and the regime
becomes supersonic. Therefore, the TR is characterized by
P′
e2=P′′
e2.(10)
By applying the mass, momentum and energy balances to the control vol-
ume defined between section e2and section e3, we can write the following
equations:
continuity equation
m′+m” = m3(11)
with
U=m”
m′.(12)
The Eq. (11) can be expressed as
m3
m′= 1 + U . (13)
Energy equation
m′CpT′
0+m”CpT0” = m3CpT0e3,(14)

Hybrid cascade refrigeration systems for refrigeration and heating 81
where T0e3is the stagnation temperature at the section e3.
By dividing the two members of the equation by m′CpT′
0, the energy
equation can be expressed as follows:
(1 + Uθ) = (1 + U)(T0e3
T′
0
),(15)
where θis defined as
θ=T0”
T′
0
.(16)
Momentum equation
me3Ve3+Pe3Se3+ ∆P Se3=m′V′
e2+P′
e2S′
e2+m”Ve2” + Pe2”Se2” (17)
where ∆P Se3expresses the frictional losses inside the mixing chamber and
is obtained from the following equation:
∆P Se3=F(ρV 2
m
2)( L
D)Se3,(18)
where Fis the friction factor. With the assumption ρV 2
m=ρe3V2
e3, Eq. (18)
can be expressed as
∆P Se3=F(L
D)me3
Ve3
2.(19)
In order to simplify the equations resolution, the dimensionless velocity M
(M=V /a∗) and the new function, f1(M), are introduced:
f1(M) = 1
M+M , (20)
where a∗is the sound speed at the nozzle throat and it is given by
a∗=pkRT∗
and after a series of transformations of the expression MV +P S we obtain
MV +P S =k+ 1
2ka∗mf1(M).(21)
By using Eqs. (19) and (21) and by dividing the momentum equation,
Eq. (17) by k+1
2ka′
∗m′, finally Eq.(17) can be expressed as
a∗e3
a′
∗
me3
m′hf1(Me3) + xMe3i=f1(M′
e2) + a∗”
a′
∗
m”
m′f1(Me2”) ,(22)

82 K. Megdouli, Y. Ezzaalouni, E. Nahdi, A. Mhimid and L. Kairouani
where xis defined as x=k
k+1
L
DF.
The distance L(Fig. 3) in the TR is assumed to be 10 D. By using the
isentropic relations, the momentum equation becomes:
1 + U(θ)1
2hf1(Me3) + xMe3i=f1(M′
e2) + U(θ)1
2f1(Me2”) .(23)
Calculation of the stagnation pressure P0e3in the section e3: The primary
mass flow rate can be expressed as
m′=P′
0
1
qT′
0
S′
∗k
R
1
22
k+ 1k
k−1k+ 1
2
1
2
.(24)
The mass flow rate in the section e3is
me3=P0e3
1
√T0e3
S∗e3k
R
1
22
k+ 1k
k−1k+ 1
2
1
2
,(25)
where S∗e3is the fictitious throat in the mixing chamber and can be ex-
pressed as follows by using isentropic relations:
S∗e3=Se3f2(k, Me3),(26)
where
f2(k, M ) = S∗
S=k+ 1
2
1
k−1
M1−k−1
k+ 1M2
1
k−1
.(27)
By using Eq. (24) attributed to the section e2and Eq. (25), we obtain
me3
m′
e2
=P0e3
P′
0T′
0
T0e3
1
2Se3
S′
∗
f2(k, Me3).(28)
Therefore
P0e3
P′
0
=1 + U(θ)1
2
Φf2(k, Me3),(29)
where Φ = Se3
S′
∗. Calculation of the static pressure Pe4in the section e4:
By using isentropic relations we can write
Pe4=P0e41−k−1
k+ 1M2k
k−1
,(30)

Hybrid cascade refrigeration systems for refrigeration and heating 83
where
f3(k, M ) = 1−k−1
k+ 1M2k
k−1
.(31)
We can also write
Pe4=P0e3ηDf3(k, Me4),(32)
where ηDis the pressure coefficient in the diffuser and is expressed by
ηD=P04
P03
.(33)
By using a relation similar to Eq. (16) the mass flow rate at the exit of the
diffuser can be expressed as
m4=P0e4
1
√T0e4
Se4k
R
1
22
k+ 1k
k−1k+ 1
2
1
2
f2(k, Me4).(34)
By combining Eqs. (25), (26), and (34) we obtain
me4
me3
=P0e4
P0e3T0e3
T0e41
2Se4
Se3
f2(k, Me4)
f2(k, Me3).(35)
The heat loss in the diffuser is neglected; the energy conservation in the
diffuser imposes
m3CpT03 =m4CpT04 .(36)
Therefore the stagnation temperatures T03 and T04 are equal. By using
Eqs. (33), and (36) in (34) we obtain
f2(k, Me4) = f2(k, Me3)
ΩηD
,(37)
where
Ω = Se4
Se3
.
Substituting Eqs. (29) and (29) in Eq. (32), we can find a relation between
the exit parameters (Pe4,Me4) and the inlet parameters (P′
0,θ):
f2(k, Me4)
f3(k, Me4)=ξ1 + U(θ)1
2
ΦΩ .(38)

84 K. Megdouli, Y. Ezzaalouni, E. Nahdi, A. Mhimid and L. Kairouani
Entrainment ratio U: Similar to Eq. (28), the entrainment ratio Ucan
be expressed as
U=m”
m′=P′′
0
P′
0T′
0
T′′
0
1
2S′′
e2
S′
∗
f2(k, M ′′
e2).(39)
Finally, by using Φ we obtain
U θ 1
2=P′′
0
P′
0Φ−1
f2(k, M ′
e2)f2(k, M ′′
e2).(40)
On the other hand, by using the hypothesis expressed by P′
e2=P′′
e2, and
similarly to Eq. (32), we have
P′
e2=P′
0f3(k, M ′
e2)
and
P′′
e2=P′′
0f3(k, M ′′
e2),
hence P′
0
P′
0
=f3(k, M ′′
e2)
f3(k, M ′
e2).(41)
By combining Eqs. (40) and (41) we get
f2(k, M ′′
e2)
f3(k, M ′′
e2)=U(θ)1
2
(Φ −1
f2(k,M ′
e2))f3(M′
e2).(42)
System of equations: The final system of equations used for the transi-
tion regime is given by
1 + U(θ)1
2hf1(Me3) + xMe3i=f1(M′
e2) + U(θ)1
2f1(M′′
e2),(43)
f2(k, Me4) = f2(k, Me3)
ΩηD
,(44)
f2(k, M e4) = f3(k, Me4)ξ(1 + U(θ)1
2)
ΦΩ ,(45)
U(θ)1
2=1
ΓΦ−1
f2(k, M ′
e2)f2(k, M ′′
e2),(46)
Γ = 2
k+ 1(k/k−1) 1
f3(k, M ′
e2).(47)

Hybrid cascade refrigeration systems for refrigeration and heating 85
In which
Γ = P′
0
P′
0P′′
0
,(48)
ξ=P′
0
P′
0Pe4
,(49)
r=Pe4
Pe4P′′
0
,(50)
or
r=Γ
Γξ.
In this system of equations, we have nine parameters: U(θ)1
2,ξ, Γ, Φ, Ω,
M′
i,Mi”, M3, and M4(by supposing that the pressure coefficient in the
diffuser ηDand the friction factor F(therefore xare fixed). The most
important parameters are thermodynamic parameters U(θ)1
2,ξ, Γ and ge-
ometric parameters: Φ, and Ω. To have the solution of the system, it is
necessary to fix four initial parameters. In our case, the four fixed variables
are: M′′
e2, Γ, Φ and Ω. The five unknown parameters are U(θ)1
2,ξ,M′
e2,
Me3, and Me4.
The aim of the ejector modeling is to find the back pressures the ejector
and the entrainment ratio.
4 Exergetic analysis
Heat transfer between a thermodynamic system and the surroundings is the
most important source of irreversibility. This irreversibility causes a degra-
dation of the system performance. Exergy analysis is a powerful tool to
locate this irreversibility. It expresses the maximum quantity of work that
would be possible for withdrawal by means of a driving thermodynamic
cycle.
Exergy losses or irreversibility are defined as the difference between the
exergy inputs to the process and the exergy output.
The total exergy destruction rate of the cycle is considered as the sum of
the exergy destruction rates in each component:
Itot =Ic+Ie+Icd +Iev1+Iint +Ib+Ip+Iej +Iev2.(51)

86 K. Megdouli, Y. Ezzaalouni, E. Nahdi, A. Mhimid and L. Kairouani
Table 1: Shows the component’s irreversibility for the cascade system with two ejectors.
Subsystems Exergy relations
Compressor Ic=T0˙m1(s2−s1)
Evaporator Ie= ˙m1T0(s4−s1+ (h4−h1)/Tr)
Intercooler Iint =T0[U(s5−s6)/(1 + U) + ˙m1(s3−s2)]
Expansion device1 Iev1=T0˙m1(s4−s3)
Expansion device2 Iev2=T0(s6−s7)U/1 + U
Condenser Icd = [h10 −h7−T0(s10 −s7)]
Ejector Iej =T0[s10 −s9/(1 + U)−s5U/1 + U]
Pump Ip=T0(s7−s8)/(1 + U)
Generator Ib=T0[s9−s8+ (h8−h9)/Tg]/(1 + U)
The second-law efficiency, ηII , is a valuable parameter to evaluate the per-
formance of cycle and can be expressed as follows:
ηII =˙
Qe(T0/Tr−1) + ˙
Qcd(1 −T0/Th)
˙
Wc+˙
Wp
.(52)
5 Validation of model results
Based on the above model a simulation program using the database REF-
PROP V 9.1 was developed. It determines the thermodynamic properties
and the mass flow rate of working fluids at all the states identified in Fig. 1.
The validation of the computer simulation of the ejector vapor-compres-
sion cycle is carried out for R134a with Tb= 78 ◦C and Tc= 40 ◦C, by
varying the intercooler temperature from 13 to 25 ◦C. The simulated perfor-
mance is compared with that of experimental data available in the literature
[10]. The simulated results agree fairly well with the experimental data,
shown in Tab. 2, and the maximum deviation is found to be 6.06%.

Hybrid cascade refrigeration systems for refrigeration and heating 87
Table 2: Comparison of the present model results with the experimentally validated re-
sults.
Te(◦C) COP Relative error (%)
Jia et al. [10] Present study
13 3.5 3.7 5.71
16 3.4 3.6 5.88
19 3.3 3.5 6.08
22 3.25 3.4 4.61
25 3.19 3.3 3.44
6 Influence of fluid nature on exergy loss and
system performance
For fixed operate conditions (Te= 10 ◦C, Tc= 80 ◦C, Tb= 65 ◦C,
Tr= 15 ◦C, T0= 25 ◦C, Th= 75 ◦C, and Tg= 70 ◦C), the temperature
difference in the intercooler between the two subcycles, ∆T, is assumed to
be 5 ◦C [13,14], 2017). Several fluids have been studied to find the fluid
which gives the maximum performance.
The refrigerants considered in this study are pure fluids. R1270, R125,
R143a, R32, R41 show higher COP than R22. We found that the COP is
improved by 3.84%, 4.7%, 4.27%, 8.97%, and 10.68% respectively (Fig. 4).
Figure 5, shows the second law efficiency for various pure fluids. The R32
gives the best performance. To locate sources of losses, we investigate the
behavior of cascade system using the exergy method. In Fig. 6, the exergy
loss of R227ea, R236ea, R236fa, R32, Rc318, R134a, R1234yf, and R1234ze
is clearly lower than for R22.
All in all and based in the energy and exergy analysis R32 still becomes
the best choice to be used in this proposed system. The aim of this work
is to study the performance of this fluid.

88 K. Megdouli, Y. Ezzaalouni, E. Nahdi, A. Mhimid and L. Kairouani
Figure 4: COP variation versus different pure fluids.
Figure 5: Second law efficiency variation versus different pure fluids.

Hybrid cascade refrigeration systems for refrigeration and heating 89
Figure 6: Total irreversibility loss variation versus different pure fluids.
7 Influence of condensation temperature on
system performance and exergy loss
Figures 7 and 8 display the effect of the condensation temperature on the
coefficient of performance and the second law efficiency of the system. The
COP and the second law efficiency decrease with increasing condensation
temperature, since the exergy loss of the overall system increases (Fig. 9).
The exergy loss of each component for the R32 are plotted in Fig. 10.
Remarkably, the exergy destruction rates of the components in the com-
pressor heat pump except the intercooler grow while the others’ tend to be
constant.
The largest loss 39.19% occurs in the condenser followed by the expan-
sion valve 8.66% for Tc= 55 ◦C. The rest of the exergy loss occurs in the
compressor, pump, evaporator, ejector, intercooler and the boiler.

90 K. Megdouli, Y. Ezzaalouni, E. Nahdi, A. Mhimid and L. Kairouani
Figure 7: COP variation versus condensation temperature.
Figure 8: Second law efficiency variation versus condensation temperature.

Hybrid cascade refrigeration systems for refrigeration and heating 91
Figure 9: Total irreversibility loss variation versus condensation temperature.
Figure 10: Effect of condensing temperature on exergy loss of each component (R32).

92 K. Megdouli, Y. Ezzaalouni, E. Nahdi, A. Mhimid and L. Kairouani
8 Influence of evaporation temperature system
performance on exergy losses
The operating conditions are chosen such as Tc= 75 ◦C, Tb= 90 ◦C, and
∆T= 5 ◦C. The temperature of evaporator is varied between −15 ◦C and
20 ◦C. It is observed that increasing the evaporation temperature increases
COP (Fig. 11) however the total exergy loss decreases (Fig. 13). The Fig. 12
indicates that the evaporator temperature has less influence on the second
law efficiency.
Figure 11: COP variation versus evaporating temperature
The exergy loss decreases in all components of the proposed cycle when
the evaporation temperature rise from −15 to 20 ◦C. The decrease in the
irreversibility in the condenser and the expansion valve (Fig. 14) decreases
the overall irreversibility and consequently higher performance.
9 Conclusion
The impact of three design parameters has been studied, namely the con-
denser temperature, evaporator temperature, and the nature of fluid. The
important conclusions which can be drawn from this study can be summa-

Hybrid cascade refrigeration systems for refrigeration and heating 93
Figure 12: Second law efficiency variation versus evaporating temperature.
Figure 13: Total irreversibility loss versus evaporating temperature.

94 K. Megdouli, Y. Ezzaalouni, E. Nahdi, A. Mhimid and L. Kairouani
Figure 14: Effect of evaporating temperature on exergy loss of each component (R32).
rized as follows:
• It is observed that the coefficient of performance and the second law
efficiency increased with the increasing evaporator temperature or
decreasing the condenser temperature.
• Exergy analysis is good for finding the losses in the system.
• It is hoped that these theoretical results will stimulate wider interest
in the technology of the application of R32 as an alternative refriger-
ant to R22 for refrigeration and heat pump applications.
Received April 2017
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